EP2084395B1 - Ensemble de joint d'étanchéité entre deux parties mobiles l'une par rapport à l'autre d'une turbomachine - Google Patents

Ensemble de joint d'étanchéité entre deux parties mobiles l'une par rapport à l'autre d'une turbomachine Download PDF

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Publication number
EP2084395B1
EP2084395B1 EP07815157A EP07815157A EP2084395B1 EP 2084395 B1 EP2084395 B1 EP 2084395B1 EP 07815157 A EP07815157 A EP 07815157A EP 07815157 A EP07815157 A EP 07815157A EP 2084395 B1 EP2084395 B1 EP 2084395B1
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EP
European Patent Office
Prior art keywords
bearing
hydraulic
sealing ring
sealing
arrangement according
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EP07815157A
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German (de)
English (en)
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EP2084395A1 (fr
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Philipp Gittler
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16J—PISTONS; CYLINDERS; SEALINGS
    • F16J15/00—Sealings
    • F16J15/44—Free-space packings
    • F16J15/443—Free-space packings provided with discharge channels
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F03—MACHINES OR ENGINES FOR LIQUIDS; WIND, SPRING, OR WEIGHT MOTORS; PRODUCING MECHANICAL POWER OR A REACTIVE PROPULSIVE THRUST, NOT OTHERWISE PROVIDED FOR
    • F03B—MACHINES OR ENGINES FOR LIQUIDS
    • F03B11/00—Parts or details not provided for in, or of interest apart from, the preceding groups, e.g. wear-protection couplings, between turbine and generator
    • F03B11/006—Sealing arrangements
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F05—INDEXING SCHEMES RELATING TO ENGINES OR PUMPS IN VARIOUS SUBCLASSES OF CLASSES F01-F04
    • F05B—INDEXING SCHEME RELATING TO WIND, SPRING, WEIGHT, INERTIA OR LIKE MOTORS, TO MACHINES OR ENGINES FOR LIQUIDS COVERED BY SUBCLASSES F03B, F03D AND F03G
    • F05B2240/00—Components
    • F05B2240/57—Seals
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F05—INDEXING SCHEMES RELATING TO ENGINES OR PUMPS IN VARIOUS SUBCLASSES OF CLASSES F01-F04
    • F05B—INDEXING SCHEME RELATING TO WIND, SPRING, WEIGHT, INERTIA OR LIKE MOTORS, TO MACHINES OR ENGINES FOR LIQUIDS COVERED BY SUBCLASSES F03B, F03D AND F03G
    • F05B2260/00—Function
    • F05B2260/60—Fluid transfer
    • F05B2260/602—Drainage
    • F05B2260/603—Drainage of leakage having past a seal
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02E—REDUCTION OF GREENHOUSE GAS [GHG] EMISSIONS, RELATED TO ENERGY GENERATION, TRANSMISSION OR DISTRIBUTION
    • Y02E10/00—Energy generation through renewable energy sources
    • Y02E10/20—Hydro energy

Definitions

  • the subject invention relates to an arrangement for sealing between two relatively moving parts of a hydraulic fluid machine, preferably between a housing and an impeller of a hydraulic fluid machine, with a sealing element which is arranged floatingly mounted with respect to the two parts on a first and a second hydraulic bearing , wherein the first and the second hydraulic bearing consist of at least two associated bearing surfaces and the sealing element is supported by acting on its bearing surfaces hydraulic pressure forces and wherein in the sealing element, a hydraulic connection, preferably a connecting bore, is provided which the first hydraulic bearing with the hydraulically connects second hydraulic bearing and with a supply line, which opens into the first hydraulic bearing.
  • the US 5 052 694 A shows a sealing body which is mounted on a hydrostatic bearing in one direction. In a second direction, this sealing body is forcibly guided by means of sealing elements.
  • a throttle device is provided, are damped with the variations in the gap widths of the hydrostatic bearing.
  • Also in the AT 298 365 B discloses a seal with a sealing body which is mounted in one direction on a hydraulic bearing, in the other Direction by a piston and a line but again positively driven. The necessary contact pressure of the sealing body is effected by the hydraulic piston.
  • the supply pressure of the hydraulic bearing and the piston is set by separate, independent of each other throttles in the respective supply line.
  • seals between two relatively moving parts by means of a hydrostatic bearing.
  • a sealing element is mounted in a direction between the two relatively moving parts on a hydrostatic bearing.
  • the sealing element is mounted on a hydraulic bearing.
  • the hydraulic bearing consists in each case of two sealing bodies, such as O-rings, between which a pressure medium is supplied. Examples of such seals are the US 4 118 040 A or the GB 839,880 refer to.
  • the sealing ring has in each case at least one connecting bore, which hydraulically connects the two hydrostatic bearings.
  • the sealing water is now fed via at least one supply line to a warehouse and passes through the connecting hole (s) in the second camp.
  • the connecting bores have always been made as large as possible and / or provided with many circumferentially distributed connecting bores in order to minimize the unavoidable natural flow losses, which, as known, depend essentially on the geometry of and the flow velocity through the connecting bores.
  • the disadvantage of the disclosed in these two applications arrangements of sealing rings in the turbine is firstly that the sealing ring in practice has unsatisfactory functionality, especially in a sealing ring according to the Fig. 3 of the WO 02/23038 A1 , If the sealing ring is radially fixed to the housing, then the sealing ring can be made difficult to lift in the radial direction, since the entire sealing water which is pushed in via the supply line is passed through the connecting bore to the second bearing and leads there to a strong lifting in axial direction. The sealing ring would thus grind on the housing, which leads to damage and can result in destruction.
  • the ring in case the ring is installed with some radial play, it would center in operation and rise radially and axially, but it would not take a favored position due to the lack of balance of forces, it would be unstable and it would not be radially as well adjustable. If one tries to change namely the radial position by changing the volume flow, only the axial position would change, since a changed volume flow would in turn be forwarded via the connecting bore directly to the axial bearing. Such a sealing ring would therefore be less practical in practice.
  • the sealing ring sealing water is required with a pressure that is significantly greater than the upper water pressure itself to lift the sealing ring and to keep it in a stable state. In the operation of the sealing ring, therefore, an additional aggregate corresponding power to increase the pressure of the sealing water, such as a water pump is required, which of course increases the effort and makes the gain in efficiency through the sealing ring partially canceled out.
  • the object of the present invention is therefore to remedy the existing with the known sealing ring arrangements still existing disadvantages, in particular to provide a compact arrangement can be dispensed with additional units to increase the sealing water pressure or much smaller additional units are needed and still a stable Operation of the sealing ring is made possible.
  • a Throttle device is provided, with which a pressure loss, based on the sealing element applied to the sealing pressure to be sealed, of at least 1%, preferably at least 5%, preferably at least 10%, between the first hydraulic bearing and second hydraulic bearing is predetermined.
  • the barrier water can eg be taken directly from the upper water and adjusted by pressure control devices to the desired barrier water pressure. This results in a significantly lower energy loss than in the known generic same sealing rings of the prior art.
  • the arrangement according to the invention also enables a stable operation of the sealing element, since it is possible by the internal throttling between the two hydraulic bearings, also to control the radial bearing.
  • the hydraulic bearings are designed as hydrostatic bearings.
  • the throttle device is designed as a hydraulic connection, pipe insert in the hydraulic connection, as a diaphragm or perforated plate or as a sudden widening of the cross section of the hydraulic connection.
  • the sealing element is arranged secured against rotation by means of an anti-rotation against one of the two parts of the hydraulic fluid machine, wherein the rotation on the sealing element advantageously substantially only exerts a circumferential force.
  • sealing rings with a large diameter it is advantageous to perform this divided into at least two parts, which simplifies the production and transport of the sealing ring.
  • the sealing ring is slotted at a number of pitch points and the remaining parting points, the parts of the sealing ring are rigidly connected to each other, wherein at the slotted division point in the circumferential direction, a gap is provided.
  • a means for preventing a flow can be provided be.
  • deformation of the slotted sealing ring may be provided at the slotted division point means for preventing deformation of the sealing ring, which does not hinder the mobility of the sealing ring in the circumferential direction.
  • the sealing element can be built torsionally soft, which advantageously reduces both the dimensions and the weight of such a sealing element.
  • connection bore Under hydraulic connection, e.g. a connection bore is referred to in the sense of this application, at least one cavity with at least two open ends, said cavity can be traversed by an arbitrary way from one end to the other end of a medium.
  • the Fig. 1 now shows a turbine 1, here a Francis turbine, with an impeller 2, which runs in a turbine housing 12.
  • the impeller 2 has a number of turbine blades 3 bounded by the ring 10 and the hub 11.
  • the impeller 2 is secured by means of a hub cap 9 and possibly by means of further fastening means, such as bolts or screws, at one end of the shaft 8 with respect to the shaft 8 against rotation.
  • the shaft 8 is rotatably supported by means not shown shaft bearings and drives in a known manner, for example, a generator, also not shown, for generating electrical energy, which is preferably arranged at the other end of the shaft 8.
  • volute casing The inflow of the liquid medium, usually water, from an upper water, such as a higher-lying water reservoir, in most cases via a not shown here, well-known volute casing.
  • a diffuser 4 consisting of a number of vanes 5, which are rotatable in this example by means of an adjusting device 6.
  • the adjustable vanes 5 serve to regulate the power of the turbine 1 by varying both the volume flow through the turbines 1 and the impeller inlet swirl.
  • the volute casing and guide vanes 5 could also be arranged in a known manner supporting blades.
  • FIG. 2 is now a detailed view of an exemplary inventive sealing of a peripheral part of an impeller 2 of a turbine 1 between the turbine housing 12 and hub 11 by means of a sealing ring 20 designed as a sealing element on two hydraulic bearings, in this case two hydrostatic bearings 22, 23, mounted floating is shown.
  • Floating means here that the sealing element is freely movable substantially in all directions and the movements of the sealing element no other forces, such as holding forces, frictional forces, forces caused by further sealing body between the sealing element and parts of the turbine 1 or forces further counteracting hydraulic bearings , counteract in the same order of magnitude as acting on the sealing element hydraulic pressure forces.
  • the turbine housing 12 a shoulder 35 on which a radial bearing surface 26 is arranged.
  • an axial bearing surface 27 is arranged on the impeller 2 on the hub 11.
  • These bearing surfaces 26, 27 may be separate components, which are subsequently applied to the required location, for example by welding, screwing, gluing, etc., or of course, directly in the corresponding component, such as the hub 11, incorporated, for example, a flat ground and with a wear-resistant coating coated portion on the hub 11 or the shoulder 35.
  • the radial bearing surface 26 on the turbine housing 12 and the axial bearing surface 27 on the hub 11 is assigned a respective radial 34 or axial bearing surface 33 on the sealing ring 20.
  • the two associated bearing surfaces 26, 34 and 27, 33 each form part of a radial and axial hydrostatic bearing 22, 23rd
  • orientations "axial” and “radial” always refer to the directions of action of the hydraulic bearings with respect to the axis of rotation of the impeller 2 of the turbine 1 and are mainly for ease of distinctness of the two hydraulic bearings, such as the two hydrostatic bearings 22, 23, introduced in the description.
  • a storage medium such as water (sealing water) fed into the radial hydrostatic bearing 22.
  • the supply line 21 is formed here from holes in the turbine housing 12, which optionally via other lines indirectly or directly to a supply source, not shown, such as a pump and / or the upper water, possibly on auxiliary equipment such as filters, cyclones, chokes, flow control valves, etc., connected is.
  • a throttle or a flow control valve in the supply line 21, it is very easy to set the supplied quantity of sealing water Q or the sealing water pressure p 1 .
  • a plurality of supply lines 21, preferably symmetrical, may be distributed over the circumference, with an arrangement favorable for a sufficient and uniform supply, e.g. three supply lines 21, which are arranged offset by an angle of 120 °, may be provided.
  • any other sensible arrangement is conceivable.
  • the radial bearing 22 now has a bearing recess in the form of a groove 24, into which the supply line 21 opens and which, as known, (idealized) forms a region of constant pressure.
  • This groove 24 of the radial hydrostatic bearing 22 is connected via a hydraulic connection, here a connecting bore 30, with a groove 25 of the axial bearing 23 arranged in the sealing ring 20.
  • the two grooves 24, 25 are also part of the associated hydrostatic bearings 22, 23rd
  • bearing recesses here grooves 24, 25 also equivalent in the axial or radial bearing surface 26, 27 of the turbine housing 12 or the impeller 2, as here in the hub 11, could be arranged. It would also be possible to provide bearing recesses both in the sealing body 20 and in the turbine housing 12 or at any point of the impeller 2 in the region of the associated hydraulic bearing. Likewise, it is conceivable that a hydraulic bearing has no bearing recess at all.
  • the LagerausEnglishept here the grooves 24, 25, of course, must not be connected continuously over the circumference, but it is also possible over the circumference several distributed, the grooves 24, 25 forming pockets to arrange, as in Fig. 3 shown and described in detail below.
  • both hydrostatic bearings 22, 23 are supplied with storage medium, the sealing water, from a single supply line 21 (or from a number of circumferentially distributed supply lines).
  • the sealing water is supplied in the present example in the radial bearing 22 and flows through the connecting holes 30 in the axial bearing 23.
  • the supply line 21 but could of course also open in the axial bearing 23.
  • a throttle device 37 which causes a predetermined pressure drop between the first hydrostatic bearing 22 and the second hydrostatic bearing 23.
  • the throttle effect of a throttle device 37 is essentially proportional to a loss coefficient ⁇ , to the density ⁇ of the flowing medium and to the square of the flow velocity through the throttle device 37, according to basic hydrodynamic laws.
  • the throttle device 37 can be designed in different ways, eg as a diaphragm, as a perforated plate As a baffle, by shaping the geometry, such as an appropriate choice of the diameter or length, the connecting hole 30, or it can, as in the example according to Fig. 2 , the known Carnot shock loss are exploited in a sudden expansion of the cross section of a flow channel.
  • Fig. 2 is in the region of the groove 24 in the input region of the connecting bore 30, a region 31 provided with a large cross-section, from which the connecting bore 30 extends, wherein the transition from the input portion 31 may be rounded to the connecting hole 30 in order to reduce possible wear in this area and thus the selected throttling effect over longer periods of the sealing element 20 to keep constant.
  • the opposite end 32 of the connecting bore 30 has a sudden cross-sectional widening, at which the Carnot shock loss arises.
  • the speed through the connecting bore 30 can, at a certain pressure of the sealing water p 1 or at a certain volume flow Q (which can be controlled for example by a flow control valve in the supply line 21), by the cross-sectional area of the connecting bore 30 and / or preferred be set by the number of distributed over the circumference arranged connecting holes 30.
  • the throttle device 37 is now chosen or designed so that between the first hydrostatic bearing 22 and the second hydrostatic bearing 23, a pressure loss, based on the pressure applied to the sealing ring 20 pressure difference p o -p u , of at least 1%, preferably at least 5 %, preferably at least 10%.
  • the natural unavoidable pressure loss caused by a flow through as lossless as possible bore, as desired in previous sealing rings is typically smaller by about two to three powers of ten, so it is negligible compared to the adjustable "internal throttling".
  • the desired pressure loss according to the invention by the throttle device 37 is significantly above the unavoidable natural throttle effect (essentially by pipe friction, possible deflections in the connection bore and due to a cross-sectional widening).
  • connection bore 30 can also be achieved by a tubular insert 38 which is arranged in a larger through bore, as in Fig. 5 shown.
  • a tubular insert 38 which is arranged in a larger through bore, as in Fig. 5 shown.
  • the storage medium is supplied via the supply line 21 with a volume flow Q or at a pressure p 1 in the radial bearing 22 as described above.
  • the volume flow Q of the storage medium is divided in the radial bearing 22 in three streams.
  • One Current flows down and finally flows with the pressure p u in the axial wheel side space.
  • a second current flows upwards and opens with the prevailing at the impeller inlet pressure p o , which of course also acts on the sealing ring 20, in the storage room 36.
  • a third, by far the largest flow flows through the connecting hole 30 in the axial bearing 23. Die
  • the distribution of the volume flow Q of course results from the radial bearing gap, which is set essentially by the pressure p 1 prevailing in the hydrostatic bearing 22 and the pressure p o at the impeller inlet.
  • the volume flow Q caused by fundamental known fluid mechanical laws and depending on the geometry of the radial bearing 22 in the radial bearing 22, the pressure distribution shown with a maximum pressure p 1 in the groove 24, in which the supply line 21 opens.
  • This pressure distribution causes the sealing ring 20 lifts in the radial direction.
  • Radial lifting means in this context, on the one hand, that the sealing ring 20 is centered accordingly, since the pressure acts over the entire inner circumference, and on the other hand widens. This widening is counteracted both by the upper water pressure p o and also by the elasticity theory, the elastic restoring forces of the sealing ring 20.
  • the maximum pressure p 1 must therefore be large enough to effect the lifting of the sealing ring 20 to the desired bearing gap, for example typically 50-100 ⁇ m.
  • a pressure p 2 which is smaller than the pressure p 1 in the radial hydrostatic bearing 22 in accordance with the adjustable pressure loss between the two bearings 22, 23.
  • a pressure distribution in the axial bearing 23 can be adjusted to lift the sealing ring 20 in the axial direction.
  • connection bore 30 between the radial hydrostatic bearing 22 and the axial hydrostatic bearing 23 can via the pressure in the supply line 21 and the fed volume flow Q are also affected the pressure p 1 in the radial bearing 22 , whereby also the radial bearing 22 can be controlled.
  • the sealing ring 20 may have large diameters in real applications, up to a few meters, but small cross sections and lightweight materials are desired, the sealing ring 20 is very torsionally soft, so very sensitive to moments, such as tilting or Stülp moments. Therefore, it is important to keep the sealing ring 20 during operation torque-free, otherwise a very rigid deformation ring would be needed.
  • the setting of a moment equilibrium is in this case by the choice of the geometry of the sealing ring 20, the hydrostatic bearing 22, 23, and the pressure distribution and the pressure level in the camps 22, 23 possible. This can, as one can easily think, be achieved by the grooves 24, 25 are arranged so that the torque balance is balanced by the area acting on the sealing ring 20 pressure distributions to the area cross section of the sealing ring cross-section. In order to achieve this, in addition to the entire geometry of the sealing ring 20, such as the position and dimensions of the grooves 24, 25, the bearing gap widths, the outer dimensions, etc., among other things, the recess 29 is used.
  • the sealing ring 20 thus floats virtually smoothly and stably on two hydraulic bearings and is held in equilibrium only by the pressure acting on the sealing ring 20 hydraulic pressure forces. In equilibrium means that the sealing ring 20 is both in the balance of power, as well as torque-free. If the sealing ring 20 is not arranged secured against rotation, it also turns at a lower speed with the impeller 2. This also results in a dynamic stability gain, since thus the limit peripheral speed or the flutter limit is set up. In addition, the lower relative velocities also reduce the friction losses.
  • the rotation can under the condition that the axial and radial mobility of the sealing ring 20 may not be obstructed in operation or only negligible, are carried out almost arbitrarily.
  • a simple stationary stop on the turbine housing 12 are provided, against which a correspondingly shaped stop on the sealing ring 20 is applied to prevent the rotation of the sealing ring 20. Since the sealing ring 20 as described above floats almost lossless and smoothly on the resulting in the hydraulic bearings pressure pads, only very small circumferential forces on the sealing ring 20, typically in the range of a few hundred N or smaller.
  • the sealing ring 20 Due to the high stability of a hydraulic bearing and the free mobility of the sealing ring 20, the sealing ring 20 is able to compensate for vibrations of the impeller 2 and / or the turbine housing 12, and axial displacements of the impeller 2, without losing the sealing effect and without Contact with the impeller 2 and / or the turbine housing 12 to get to.
  • the sealing ring 20 suffers practically no wear because it floats virtually smoothly on liquid films, whereby the life of such a sealing ring 20 is very high. Due to the fact that the sealing ring 20 can preferably be constructed as a very slender, lighter ring, which has hardly any mass forces, this effect is enhanced.
  • the sealing ring 20 can be made very small in relation to the dimensions of the impeller 2, edge lengths of a few centimeters, e.g. 5cm or 8cm, with outside diameters of a few meters are quite sufficient, and it can be made of any material such as steel, bearing bronze, plastic (e.g., PE). Furthermore, the bearing surfaces 26, 34 and 27, 33 could also be coated with a suitable layer, such as Teflon, bearing bronze, etc., in order to further improve the properties of the seal. Typically, the sealing ring 20 will be made of a softer material, such as e.g. Teflon, bronze, etc., made as the housing 12 or the impeller 2 of the hydraulic machine. As a result, he is on the one hand usually easier and on the other hand, in extreme cases, the sealing ring 20 and not the impeller 2 or the housing 12 is damaged or even destroyed.
  • a softer material such as e.g. Teflon, bronze, etc.
  • the Fig. 3 shows a perspective view of a sealing ring 20 according to the invention on one end face of the sealing ring 20, a plurality of grooves 25 are incorporated, which are part of the axial hydrostatic bearing 23, wherein in each of these grooves 25 at least one connecting bore 30 opens.
  • a series of grooves 24 is arranged, which are part of the radial hydrostatic bearing 22 and wherein in each of these grooves 24 in turn at least one connecting bore 30 opens.
  • the number, size and distribution of the grooves 24, 25 and the connecting bores 30 can be selected accordingly to achieve a balance of forces and moments and to set the desired pressure distributions and the bearing gaps in the associated hydrostatic bearings 22, 23.
  • the sealing ring 20 can also, as in Fig. 3 and 4 shown at one or more dividing points 41, 44 executed divided in the circumferential direction, which facilitates, among other things, the production or the transport of the sealing ring 20.
  • the sealing ring 20 is divided in the embodiment shown at two graduation points 41, 44, wherein the two halves are rigidly connected to one another at a first division point 41 via a connecting means, such as via dowel pins and a screw.
  • a connecting means such as via dowel pins and a screw.
  • the sealing ring 20 remains slotted in the assembled state, so that a gap 50 is formed in the circumferential direction, as in FIG Fig. 4 shown.
  • This gap 50 serves essentially to allow an automatic circumferential length compensation, the possible wear of the sealing ring 20 on the bearing surface 34, for example by fine solids (such as sand) in the supplied Sealing water, which have an abrasive effect, compensates.
  • the automatic circumferential length compensation results from the fact that the bearing gaps (ie, the thickness of the liquid films in the bearings) are adjusted by the supply of a certain flow with a certain pressure.
  • the bearing gaps ie, the thickness of the liquid films in the bearings
  • the slotted sealing ring 20 is compressed in the radial direction (the gap 50 at the division point 44 is thus smaller), so that the desired bearing gap in the hydrostatic bearing sets again. If the entire gap 50 is "used up", so the two parts of the sealing ring thus abut each other and thus further wear can not be compensated, the sealing ring 20 can be exchanged.
  • a shoulder 45 is now provided on the one hand at the slotted division point 44 at a circumferentially arranged end face of the sealing ring, against which a recess 47 abuts against the opposite end face of the sealing ring 20 arranged in the circumferential direction.
  • a plate which extends over the gap 50 can be arranged on the outer end face of the sealing ring 20 facing away from a hydrostatic bearing. This would again be created a seal that prevents axial flow of sealing water through the slot. Furthermore, this would prevent axial bending and twisting of the sealing ring 20.
  • the plate could be fastened by means of suitable connecting means, such as by means of dowels and screws on the sealing ring 20, said connecting means may of course only be arranged on one half of the sealing ring 20 to the mobility of the two sealing ring halves in the circumferential direction (which yes for the handledslynausmaschine important) does not hinder.
  • the sealing ring 20 can be any other suitable arrangements in order to realize a corresponding seal and to prevent deformation of the sealing ring 20.
  • the sealing ring 20 instead of the plate, the sealing ring 20, of course, analogously to the radial direction as described above, in the axial direction on one half have a shoulder which cooperates with a corresponding recess on the other half.
  • both the radial and the axial groove 24, 25 are interrupted or the symmetrical arrangement of the grooves 24, 25 are disturbed by the circumference. This could result in lower bearing forces in the region of this division, since in this area the surfaces of the grooves 24, 25, to which the bearing pressure p 1 , p4 attacks, would be smaller, which would lead to an uneven pressure load of the sealing ring 20.
  • the grooves 24, 25 adjoining the gap 50 can be made larger, or a different groove pitch in the circumferential direction and thus again a different groove area could be provided in this area.
  • the grooves 25 adjoining the gap 50 may have a region 46 with a larger groove area on the end face of the sealing ring 20 in the end facing the gap 50.
  • the grooves 24 adjoining the gap 50 may have an enlarged groove surface 42 on the inside of the sealing ring 20 in the end facing the gap 50. With these enlarged groove surfaces 42, 46, a corresponding compensation and a restoration or maintenance of the forces and moment equilibrium can be performed.
  • sealing ring 20 is carried out co-rotating, it is advantageous, especially in a slotted or split design to balance this.
  • the sealing ring 20 may, of course, be of any cross-section, e.g. an L-shaped cross section, wherein for manufacturing considerations, a square or rectangular shape is preferred.
  • one or more of the bearing surfaces 26, 34 and 27, 33 could also be provided with well-known hydrodynamic lubrication pockets, to form an additional hydrodynamic bearing.
  • a seal according to the invention with a sealing ring 20 can of course be provided at any suitable location and is not on the embodiments according to Fig. 2 limited.
  • the sealing ring 20 could also be arranged between the end face of the impeller 2 or hub 11 and the turbine housing 12.
  • the auxiliary sealing body would of course be used so that they unfold with still sufficient sealing effect as little force as possible, since the sealing ring 20 should remain so floating. Since the sealing ring 20 is stably mounted floating on the pressure pads, this can be done easily on the Schwarzstelle height of the sealing gap. Due to the possible low force on the auxiliary sealing body, typically smaller by a factor of 1000 or more than the hydraulic bearing forces, one can again speak of a floating bearing and it can be ensured that the axial and radial mobility of the sealing ring 20 is only insignificantly and negligibly affected ,
  • the seal described above is a largely tight seal.
  • the splitting water losses thereby reduce exclusively to the escaping storage medium, so are due to the low gap heights in the hydraulic bearings, e.g. typically in the 10-60 micron range in hydrostatic bearings 22, 23, very low and can be recovered in part by introducing the cracked water into the main water flow F again.
  • water is described as the storage medium for the sake of simplicity.
  • the storage medium especially in pumps, can also be any other suitable medium, e.g. Oil, be.

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Magnetic Bearings And Hydrostatic Bearings (AREA)
  • Hydraulic Turbines (AREA)
  • Hydraulic Motors (AREA)

Claims (15)

  1. Ensemble destiné à assurer l'étanchéité entre deux parties mobiles l'une par rapport à l'autre d'une turbomachine hydraulique, de préférence entre un carter (12) et un rotor (2) d'une turbomachine hydraulique (1), avec un élément d'étanchéité (20), qui, par rapport aux deux parties, est monté flottant sur un premier et un deuxième palier hydraulique (22, 23), le premier et le deuxième palier hydraulique (22, 23) étant constitués par au moins deux surfaces d'appui (26, 34 et 27, 33) associées et l'élément d'étanchéité (20) étant monté dans deux directions différentes par les forces de pression hydraulique agissant sur ses surfaces d'appui (33, 34), et une liaison hydraulique (30), de préférence un alésage de liaison, étant prévue dans l'élément d'étanchéité (20), laquelle met en communication hydraulique le premier palier hydraulique (22) et le deuxième palier hydraulique (23), et avec une conduite d'alimentation (21), qui débouche dans le premier palier hydraulique (22) et alimente le premier palier hydraulique (22) en milieu de palier, qui s'écoule, par la liaison hydraulique (30), vers le deuxième palier hydraulique (23), caractérisé par le fait que, pour régler une perte de pression entre le premier palier hydraulique (22) et le deuxième palier hydraulique (23), un dispositif d'étranglement (37) est prévu dans la liaison hydraulique (30), une perte de pression, relativement à la différence de pression à étancher (p0-pu) qui règne au niveau de l'élément d'étanchéité (20), d'au moins 1%, de préférence d'au moins 5%, de façon plus préférée d'au moins 10%, entre le premier palier hydraulique (22) et le deuxième palier hydraulique (23).
  2. Ensemble selon la revendication 1, caractérisé par le fait que la liaison hydraulique (30) tient elle-même lieu de dispositif d'étranglement (37).
  3. Ensemble selon la revendication 1, caractérisé par le fait qu'un insert tubulaire (38) dans la liaison hydraulique (30) tient lieu de dispositif d'étranglement (37).
  4. Ensemble selon l'une des revendications 1 à 3, caractérisé par le fait que le dispositif d'étranglement (37) est réalisé sous la forme d'un diaphragme, d'une plaque perforée ou d'une chicane.
  5. Ensemble selon l'une des revendications 1 à 4, caractérisé par le fait que le dispositif d'étranglement (37) est réalisé sous la forme d'un élargissement brusque de la section transversale de la liaison hydraulique (30).
  6. Ensemble selon l'une des revendications 1 à 5, caractérisé par le fait qu'une bague d'étanchéité tient lieu d'élément d'étanchéité (20).
  7. Ensemble selon la revendication 6, caractérisé par le fait que la bague d'étanchéité (20) est réalisée séparée en au moins deux parties.
  8. Ensemble selon la revendication 7, caractérisé par le fait que la bague d'étanchéité (20) est réalisée entaillée au niveau d'un nombre de sites de séparation (44) et les parties de la bague d'étanchéité (20) sont reliées rigidement entre elles au niveau des sites de séparation restants (41).
  9. Ensemble selon la revendication 8, caractérisé par le fait qu'il est prévu, au niveau du ou des sites de séparation entaillés (44), une fente (50) permettant une compensation de la longueur de la circonférence.
  10. Ensemble selon l'une des revendications 8 ou 9, caractérisé par le fait qu' il est prévu, au niveau des sites de séparation entaillés (44), un moyen pour empêcher l'écoulement d'eau de barrage à travers les sites de séparation (44), qui ne gêne pas la mobilité de la bague d'étanchéité (20) dans le sens de la circonférence.
  11. Ensemble selon l' une des revendications 8 à 10, caractérisé par le fait qu'il prévu, au niveau des sites de séparation entaillés (44), un moyen pour empêcher une déformation de la bague d'étanchéité (20), qui ne gêne pas la mobilité de la bague d'étanchéité (20) dans le sens de la circonférence.
  12. Ensemble selon l' une des revendications 7 à 11, caractérisé par le fait que la bague d'étanchéité (20) a des évidements de palier (24, 25), et au moins un évidement de palier (24, 25) voisin d'un site de séparation (41, 44) est réalisé dans la zone du site de séparation (41, 44) avec une surface d'évidement de palier différente des autres évidements de palier ou de l'évidement de palier restant.
  13. Ensemble selon l'une des revendication 1 à 12, caractérisé par le fait que l'élément d'étanchéité (20) est disposé bloqué en rotation, au moyen d'un dispositif de blocage de rotation, par rapport à l'une des deux parties de la turbomachine hydraulique (1).
  14. Ensemble selon la revendication 13, caractérisé par le fait que le dispositif de blocage de rotation n'exerce sensiblement qu'une force périphérique sur l'élément d'étanchéité (20).
  15. Ensemble selon l'une des revendications 1 à 14, caractérisé par le fait qu'au moins un palier hydraulique (22, 23) est un palier hydrostatique.
EP07815157A 2006-11-03 2007-10-22 Ensemble de joint d'étanchéité entre deux parties mobiles l'une par rapport à l'autre d'une turbomachine Not-in-force EP2084395B1 (fr)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
AT0183706A AT504394B1 (de) 2006-11-03 2006-11-03 Anordnung zur abdichtung zwischen zwei relativ zueinander bewegbaren teilen einer hydraulischen strömungsmaschine
PCT/AT2007/000490 WO2008052231A1 (fr) 2006-11-03 2007-10-22 Ensemble de joint d'étanchéité entre deux parties mobiles l'une par rapport à l'autre d'une turbomachine

Publications (2)

Publication Number Publication Date
EP2084395A1 EP2084395A1 (fr) 2009-08-05
EP2084395B1 true EP2084395B1 (fr) 2010-09-15

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Application Number Title Priority Date Filing Date
EP07815157A Not-in-force EP2084395B1 (fr) 2006-11-03 2007-10-22 Ensemble de joint d'étanchéité entre deux parties mobiles l'une par rapport à l'autre d'une turbomachine

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Country Link
EP (1) EP2084395B1 (fr)
AT (2) AT504394B1 (fr)
CA (1) CA2668482C (fr)
DE (1) DE502007005085D1 (fr)
WO (1) WO2008052231A1 (fr)

Families Citing this family (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN101825179A (zh) * 2010-04-30 2010-09-08 天津市天发重型水电设备制造有限公司 一种混流式水轮机主轴组合密封装置
JP6428916B2 (ja) 2015-03-16 2018-11-28 Nok株式会社 シールリング
US11378188B2 (en) 2017-07-10 2022-07-05 Siemens Energy, Inc. Generator seal assembly
DE102020203767B4 (de) * 2020-03-24 2022-05-05 Eagleburgmann Germany Gmbh & Co. Kg Selbstansaugende Gleitringdichtungsanordnung
DE102020122601A1 (de) 2020-08-28 2022-03-03 Rolls-Royce Deutschland Ltd & Co Kg Dichtungssystem, Getriebe mit einem Dichtungssystem und Gasturbinentriebwerk mit einem Dichtungssystem

Family Cites Families (7)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
AT298365B (de) * 1970-06-22 1972-05-10 Andritz Ag Maschf Einrichtung zur gegenseitigen Abdichtung zweier Räume
CH598514A5 (fr) * 1975-08-29 1978-04-28 Escher Wyss Ag
US5052694A (en) * 1986-07-08 1991-10-01 Eg&G Sealol, Inc. Hydrostatic face seal and bearing
US5558341A (en) * 1995-01-11 1996-09-24 Stein Seal Company Seal for sealing an incompressible fluid between a relatively stationary seal and a movable member
WO1999027281A1 (fr) * 1997-11-21 1999-06-03 Nippon Pillar Packing Co., Ltd. Joint a gaz sans contact a pression statique
AT411092B (de) * 2000-09-15 2003-09-25 Gittler Philipp Dipl Ing Dr Te Abdichtung des laufrades von hydraulischen turbomaschinen
AT413049B (de) * 2002-07-31 2005-10-15 Philipp Dipl Ing Dr Te Gittler Dichtung zwischen zwei relativ zueinander bewegbaren teilen einer hydraulischen maschine

Also Published As

Publication number Publication date
CA2668482A1 (fr) 2008-05-08
DE502007005085D1 (de) 2010-10-28
ATE481568T1 (de) 2010-10-15
WO2008052231A1 (fr) 2008-05-08
AT504394B1 (de) 2008-10-15
CA2668482C (fr) 2015-04-28
AT504394A1 (de) 2008-05-15
EP2084395A1 (fr) 2009-08-05

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